Rotary screw compressor having a pressure bearing arrangement

ABSTRACT

A rotary screw compressor that includes a male rotor and a female rotor cooperating together within a casing. The casing has a discharge outlet connected to an outlet port at a high pressure end of the working space and a suction inlet at the low pressure end of a working space. The rotor is rotatably supported at one end thereof through a bearing arrangement that includes a bracket fixed to an end cover. The bearing bracket projects into an axial cavity provided in the rotor to form a first chamber between the bracket and the rotor. The bracket is provided with an oil feed channel to feed oil into the first chamber. The rotor is rotatably supported at the low pressure end thereof through the bearing arrangement. The corresponding bearing bracket is mounted at the low pressure end of the working space and the outer circumferential surface thereof is provided with at least a groove connected to the oil feed channel and a recess connected to an oil drainage channel provided in the bearing bracket. A seal is provided between the first chamber and the working space of the compressor.

This application is a Rule 371 continuation of PCT/NL 93/00150 filedJul. 13, 1993.

The present invention relates to a rotary screw compressor comprising acasing, a male rotor and a female rotor cooperating therewith enclosedin a working space defined by the casing, the casing having a dischargeoutlet connected to an outlet port at the high pressure end of theworking space and a suction inlet at the low pressure end of the workingspace, at least one rotor being rotatably supported at an end thereofthrough a bearing arrangement comprising a bearing bracket being fixedto an end cover and having a substantially cylindrical outercircumferential surface, the bearing bracket projecting into an axialcavity provided in the rotor forming a first chamber between the bracketand the rotor, the bracket being provided with an oil feed channel tofeed oil into the first chamber.

A rotary screw compressor oft his kind for the compression of gas isknown from JP-A-59-168290. During operation of a screw compressor therotors are subjected to radial loads arising from the compression of thegas. At the high pressure end of the working space of the knowncompressor a cylindrical bearing bracket is provided for each rotor,each bearing bracket projecting from the end cover into an internalaxial cavity provided in the high pressure end of the correspondingrotor. Pressurized oil is fed through an oil feed channel into thechamber between the bearing bracket and the rotor. The oil then leavesthe chamber and enters the working space of the compressor. Finally theoil is seperated from the compressed gas and fed into the chamber again.The rotors will also be exposed to a higher pressure at their highpressure end than at their low pressure end, resulting in an axial forceacting on each rotor towards the low pressure end. Therefore each rotorof the known compressor is provided with a rolling contact thrustbearing at the low pressure end.

The bearing arrangement of the known compressor has the disadvantagethat it has a limited load bearing capacity, particularly in the radialdirection of the rotors. Therefore the known compressor is not capableof producing a high discharge pressure or large differential pressurebetween the discharge outlet and suction inlet.

It is an object of the present invention to provide a rotary screwcompressor according to the preamble which has an improved bearingarrangement with a high load bearing capacity in order to handle a highdischarge pressure or a high differential pressure.

This has according to the invention been achieved in that at least onerotor is rotatably supported at the low pressure end thereof through thebearing arrangement, the corresponding bearing bracket being mounted atthe low pressure end of the working space and the outer circumferentialsurface thereof being provided with at least a groove connected to theoil feed channel and a recess connected to an oil drainage channelprovided in the bearing bracket, and in that sealing means are providedbetween the first chamber and the working space of the compressor.According to the invention an uncomplicated bearing arrangement isobtained capable of supporting high radial loads. The load bearingcapacities of this bearing arrangement not only arise from thehydrostatic pressure of the pressurized oil fed into the first chamberbut also from hydrodynamic load bearing effects between each stationarybearing bracket and the corresponding rotor, which will rotate at a highspeed. As the pressurized oil can also be present in the space of thefirst chamber between the end face of a bearing bracket and the bottomof the internal cavity of the rotor axial loads on the rotor can also besupported.

In a preferred embodiment the end face at the low pressure end of arotor, the end cover, the casing, and the corresponding bearing bracketdefine a second chamber, the second chamber being connected to an oilfeed channel. In this manner the pressure of the oil fed into thissecond chamber acts as a hydrostatic thrust bearing capable ofsupporting at least a part of the axial load on that rotor.

In another preferred embodiment the outer circumferential surface of atleast one of the bearing brackets is provided with two longitudinalgrooves and one recess, the recess being located on the side of thebearing bracket radially opposite the outlet port and being connected tothe oil drainage channel, the longitudinal grooves being located ateither side of the recess and being connected to the oil feed channel.The presence of two longitudinal grooves, each groove being connected tothe oil feed channel, provides a zone in the first chamber wherein ahigh oil pressure is maintained for counteracting the radial load on therotor. The location of the recess, which is connected to an oil drainchannel, on the bearing bracket radially opposite the outlet port of theworking space is preferred as an optimal counterbalancing of the radialload on the rotor can be obtained in this manner.

In a particularly advantageous embodiment the edges of the longitudinalgrooves adjacent the recess are situated in a common plane through theaxis of the bearing bracket at an equal distance from the recess, andthe edges of the longitudinal grooves most distant from the recess areeach situated in a plane inclined at an angle α to the common plane.

Preferably each recess has an approximate maximum length of 0.7 timesthe length of the bearing bracket. As each recess is located at theportion of the bearing bracket adjacent the end face thereof, a portionof the bearing bracket having a cylindrical cross section at the lowpressure side of that recess forms a restriction between the recess andthe second chamber provided at the low pressure end of the rotor. Therestriction thus obtained prevents pressurized oil from flowing from thesecond chamber towards the recess and therefore prevents a drop in oilpressure in the second chamber.

Since the radial load on the male rotor arising from the compression ofthe gas is less than the radial load on the female rotor, due to thegeometry of the rotors, the length of the bearing bracket of the malerotor and/or the length of the recess thereof is preferably less thanthe length of the bearing bracket of the female rotor and/or the recessthereof.

In another preferred embodiment a groove connected to the oil feedchannel on the bearing bracket of the male rotor and a recess on thebearing bracket of the female rotor terminate at the end face of thecorresponding bearing, and each recess on the bearing bracket of themale rotor and each groove on the bearing bracket of the female rotorare located spaced from the end face of the corresponding bearingbracket. Due to the geometry of the rotors the axial load on the malerotor arising from the compression of the gas is as a rule greater thanthe axial load on the female rotor. To compensate for this difference anadditional axial force is exerted on the male rotor as the pressurizedoil supplied to a longitudinal groove on the bearing bracket of the malerotor enters the space between the end face of that bearing bracket andthe bottom of the internal cavity of the male rotor. The return flow ofoil to the recess is obstructed and the oil pressure in this space ismaintained.

For a high-speed screw compressor capable of a high pressure differencebetween the discharge outlet and the suction inlet it is advantageousthat at least one of the rotors is provided with a ring shoulderprotruding from its low pressure end, the sealing means being providedbetween the ring shoulder and the casing. This provides a furtherincrease of the axial thrust load bearing capacity of the bearingarrangement according to the invention.

for a low-speed screw compressor with a relatively low pressuredifference and wherein cooling is obtained by feeding oil into theworking space of the compressor it is advantageous that at least one ofthe rotors is provided with sealing means between the rotor and thecorresponding bearing bracket. The low-speed screw compressor is alsopreferably provided with a rolling contact bearing between at least oneof the rotors and the corresponding bearing bracket.

The rotary screw compressor according to the present invention iscapable of achieving considerably higher differential pressures betweenthe discharge outlet and the suction inlet and considerably higherdischarge pressures than the known compressors of this kind. Traditionalscrew compressors having bearings located outside the helical screw partof the rotors are known to achieve a differential pressure of up to15-20 bar. The rotary screw compressor according to the invention canachieve high differential pressures and discharge pressures as much as 3to 4 times higher. Therefore the inventive compressor can compete withcentrifugal and piston compressors, and can be used, for example, forcompression of natural gas in gas and oil fields, in gas delivery, gasfilling and gas lift stations for gas and oil production,transportation, refinery and power recovery and chemical plants as well.Further advantages of the rotary screw compressor according to theinvention are its simple design, reliability and long service life, inparticular regarding the design of the bearing arrangements at the lowpressure end, its limited weight and small dimensions.

The invention will now be explained in greater detail through thefollowing description of preferred embodiments of the screw compressoraccording to the invention, wherein reference is made to theaccompanying drawings, in which:

FIG. 1 is a longitudinal section through the male rotor of a firstembodiment of the screw compressor according to the invention,

FIG. 2 is a section taken along line II--II of FIG. 1,

FIG. 3 is a section taken along line III--III of FIG. 2,

FIG. 4 is cross section of the bearing bracket of the male rotor of FIG.1,

FIG. 5 is a view corresponding to FIG. 2 of a second embodiment of thescrew compressor according to the invention,

FIG. 6 is a diagrammatic view, partly sectional, of a third embodimentof the screw compressor according to the invention,

FIG. 7 is a view corresponding to FIG. 6 of a fourth embodiment of thescrew compressor according to the invention, and

FIG. 8 is a view corresponding to FIG. 6 of a fifth embodiment of thescrew compressor according to the invention.

In FIGS. 1, 2 and 3 a rotary screw compressor is shown comprising acasing 1, a male rotor 6 and a female rotor 18 cooperating therewithenclosed in a working space defined by the casing. The casing has aoutlet port 2 and a discharge pipe 4 at the high pressure end of theworking space and a suction pipe 3 at the low pressure end of theworking space. Arrow A indicates the direction of the gas to becompressed. Arrow B indicates the direction of the discharge of thecompressed gas. Arrow ω indicates the rotation of the male rotor 6 whichcan be driven through drive means not shown in the drawings.

The male rotor 6 is rotatably supported through a bearing 10 at its highpressure end and a bearing bracket 11 at its low pressure end. Thebearing bracket 11 is fixed on a detachable end cover 5 of the casing 1and projects into an internal cavity in the low pressure end of the malerotor 6, thereby forming a first chamber 9 therebetween.

As can be seen in FIG. 1 the cavity and the bearing bracket 11 insidethe cavity extend over a significant part of the length of the malerotor 6. Therefore the distance between the bearings 10, 11 at oppositeends of the rotor 6 is comparatively small, as a result of which theradial forces on the rotor can be better supported through the bearingsand only a small radial deflection of the rotor will occur.

The low pressure end face of the male rotor 6 is provided with aprotruding ring shoulder 15 having a cylindrical outer surface 16. Asealing means 7 between the male rotor 6 and the casing is provided atthe high pressure end and a sealing means 8 is provided between theshoulder 15 and the casing 1 at the low pressure end.

The bearing bracket 11 has a substantially cylindrical circumferentialouter surface, the surface being provided with two longitudinal grooves25, extending parallel to the longitudinal axis of the bearing bracket,and with a recess 13. The recess 13 is an essentially rectangular cutoutformed at a distance from the substantially circular end face of thebearing bracket 11 and is connected to an oil drainage channel 12through an opening 14. As can been seen in FIG. 2 the recess 13 islocated on the side of the bearing bracket 11 radially opposite theoutlet port 2 for reasons explained further below. The longitudinalgrooves 25 are located at either side of the recess 13 seen incircumferential direction. Each longitudinal groove 25 is connected toan oil feed channel 27 provided in the bearing bracket 11 through anumber of openings 29 uniformly distributed along the length of eachgroove. As can be seen in FIG. 3 the longitudinal grooves 25 terminateat the end face of the bearing bracket 11 to provide communicationbetween each groove 25 and the space formed between the end face of thebearing bracket and the bottom of the cavity in the male rotor 6.

At the low pressure end of the male rotor 6 a second chamber 17 isformed by the annular end face of the ring shoulder 15, sealing means 8,the bearing bracket 11 and the end cover 5. The chamber 17 is connectedto oil feed channels 27 through openings 35.

The female rotor 18 is at its low pressure end rotatably supported in amanner similar to the male rotor 6. A bearing bracket 20 projects intoan internal cavity provided in the rotor 18 forming a first chamber 19therebetween. The bearing bracket 20 is mounted on the end cover 5. Thesubstantially cylindrical outer surface of the bearing bracket 20 isprovided with a recess 22 and two longitudinal grooves 24 located ateither side of the recess 22. The recess 22 is connected to an oildrainage channel 21 through an opening 23. The recess 22 is anessentially rectangular cutout and terminates at the end face of thebearing bracket 22. The longitudinal grooves 24 are located at adistance from the end face of the bearing bracket 20 and extend towardsthe low pressure end. Each longitudinal groove 24 is connected to an oilfeed channel 26 through a number of openings 28 uniformly disposed alongthe length of the groove.

The low pressure end of the female rotor 18 is provided with aprotruding ring shoulder 31 having a cylindrical outer surface 32. Asealing means 30 is provided between the shoulder 31 and the end cover 5at the low pressure end of the female rotor 18.

At the low pressure end of the female rotor 18 a second chamber 33 isformed by the annular end face of the ring shoulder 31 of the rotor,sealing means 30, the bearing bracket 20 and the end cover 5. Thechamber 33 is connected to oil feed channels 26 through openings 34.

The length of the bearing bracket 11 of the male rotor 6 projecting intothe male rotor is less than the length of the bearing bracket 20 of thefemale rotor 18 projecting into the female rotor. This is indicated bythe distance "1" in FIG. 3. Also, the length of the recess 13 is lessthan that of recess 22, both recesses having an approximate maximumlength of 0.7 times the length of the corresponding bearing bracket.

FIG. 4 shows a cross section of the bearing bracket 11 of the male rotor6. As can be seen the recess 13 is essentially a flat portion formed onthe cylindrical outer circumferential surface of the bearing bracket 11.The recess 13 communicates with the central oil drainage channel 12through the opening 14. Each groove 25 is connected to an oil feedchannel 27 through a number of openings 29 to reduce the flow resistanceof the oil feed. The longitudinal grooves 25 at either side of therecess 13 are formed such that their side edges adjacent the recess 13are located in a common first plane passing through the longitudinalaxis of the bearing bracket 11 and at an equal distance from the recess13. The other longitudinal edges of the grooves 25 are each located in asecond and third plane through the axis of the bearing bracketrespectively. The second and third plane each being inclined at an angleα, preferably equal or less than 45°, to the first plane. Thisembodiment of the bearing bracket provides optimal conditions for acombination of hydrodynamic and hydrostatic radial load bearingcapabilities and an excellent radial stiffness of the bearingarrangement. The bearing bracket 20 of the female rotor 18 has a crosssection substantially similar to that of the bearing bracket 11 of themale rotor. In an alternative embodiment not shown in the drawings thelocation of the oil feed grooves at either side of the recess on thebearing bracket can be adapted e.g. for supporting a lower radial loadon the corresponding rotor. In this case the grooves could be locatedcloser to each other, therefore a smaller zone in the first having ahigh oil pressure is obtained.

A second embodiment of the compressor according to the invention isshown in FIG. 5. The compressor is provided with bearing brackets 11, 20for the male rotor 6' and female rotor 18' respectively, the bearingbrackets being similar to the bearing brackets described hereinbefore. Asealing means 56 is provided between the bearing bracket 11 and the malerotor 6'. Towards the low pressure end of the compressor a rollingcontact bearing 57, such as a ball bearing, is mounted between the malerotor 6' and the bearing bracket 11. A sealing means 58 is providedbetween the bearing bracket 20 and the female rotor 18'. Towards the lowpressure end of the compressor a rolling contact bearing 59, such as aball bearing, is mounted between the female rotor 18' and the bearingbracket 20. This embodiment is particularly advantageous for screwcompressors operating with cooling oil injected into the gas to becompressed in the working space of the compressor. These screwcompressors operate at low speed compared with oil-free ("dry")compressors and have small clearances between the rotor teeth, andbetween the rotors and the casing. Therefore rolling contact bearings ingeneral having smaller clearances than bearing brackets are preferred.The sealing means 56, 58 can be provided in the form of a flowobstruction having a smaller clearance than the clearance between therotor and the bearing bracket. As can be seen in FIG. 4 no sealing meansare provided between the second chambers 60, 61 and the working space.

In the embodiment shown in FIG. 6 the bearing brackets 11 and 20 of themale and female rotor respectively have their oil feed channels 26, 27connected to a common source 38, e.g. an oil pump, for supplyingpressurized oil as indicated by arrow k. The oil drainage channels 12,21 of the respective bearing brackets 11, 20 are connected to an oilcollector 39. The collector 39 is vented to the atmosphere as indicatedby the arrow M. In this embodiment the source 38 is designed to supplythe oil at a pressure approximately equal to the pressure of the gas tobe compressed. This embodiment is preferred for screw compressorswherein the compressed gas has to be free of oil. Since the pressure inthe chambers 17, 33 (FIG. 3) approximates the pressure in the suctionpipe 3 the loads on the sealing means 8, 30 are limited. As the oildrainage channels 12, 21 are in open communication with the atmospherethe oil collector 39 can be of a simple design.

In the embodiment shown in FIG. 7 the bearing brackets 11 and 20 of themale and female rotor respectively have their oil feed channels 26, 27connected to a source 38 for supplying pressurized oil as indicated byarrow k. The oil drainage channels 12, 21 of the respective bearingbrackets 11, 20 are connected to an oil collector 40. The collector 40is connected to the suction pipe 3 to maintain a pressure in thecollector 40 equal to the pressure of the gas to be compressed.

In the embodiment shown in FIG. 8 the bearing brackets 11 and 20 of themale and female rotor respectively have their oil feed channels 26, 27connected to an oil separator 41 for supplying pressurized oil asindicated by arrow m. The oil drainage channels 12, 21 of the respectivebearing brackets 11, 20 are connected to the suction pipe 3 of thecompressor as indicated by arrow n. The oil will then pass through thecompressor along with the gas to be compressed resulting in a cooling ofthe gas during compression. The discharge pipe 4 of the compressor isconnected to the oil seperator 41 where the oil and the compressed gasare separated. This embodiment of the compressor is preferred if thepresence of oil in the compressed gas is allowed.

The rotary screw compressor according to the invention operates asfollows.

The gas to be compressed enters the suction pipe 3 (FIG. 1). The malerotor 6 is rotated at a speed ω by means of an external drive acting onthe male rotor 6. The gas to be compressed is entrained and compressedin chambers limited by the rotor teeth and the casing. During thecompression of the gas a force F, resulting from the differentialpressure between the discharge pipe 4 and the suction pipe 3, acts onthe rotors as is indicated in FIG. 2. This force F is composed of radialforces F₁, F₂ and axial forces F₃, F₄ acting on the rotors 6 and 18.These forces must be supported by the bearing arrangements of therotors.

To counteract these forces F₁ -F₄ pressurized oil is fed through the oilfeed channels 26, 27 (arrows D and H in FIG. 3), the openings 28, 29,and the longitudinal grooves 24, 25 of the bearing brackets 11, 20 andenters the chambers 9, 19 between each bearing bracket and thecorresponding rotor. The pressurized oil is drained from chamber 9, 19through the recess 13, 22 provided on the bearing bracket, each recessbeing connected to an oil drainage channel 12, 21 by an opening 14, 23(arrows K and E in FIG. 3).

The maximum length of the recesses 13, 22, which is approximately 0.7times the length of the corresponding bearing bracket, is preferred inthis embodiment as there must be a cylindrical section of the bearingbracket having sufficient dimensions present inside the cylindricalcavity in each rotor near the low pressure end thereof to provide arestriction between the chamber 17, 33 and the recess 13, 22,respectively.

The presence of pressurized oil in the first chambers between the rotorsand the bearing brackets gives rise to radial lifting forces F₅ and F₆(FIG. 2) acting on the rotors 6, 18 respectively. The position of eachrecess on the bearing bracket, radially opposite the outlet port 2, asshown in FIG. 2, facilitates obtaining a balance between the forces F₅,F₆ and the forces F₁, F₂. As a result of the location of thelongitudinal grooves 24, 25 a pressure zone is obtained, the pressuredifference in this zone being equal to the pressure difference betweenthe oil feed channels and the oil drainage channels.

The dimensions of the recesses 13, 22, the location and dimensions ofthe longitudinal grooves 24, 25, and the pressure levels in the oil feedchannels as well as in the oil drainage channels depend on the desiredcharacteristics of the rotary screw compressor. They are chosen suchthat the forces F₅ and F₆ compensate the major part of the forces F₁, F₂respectively. The remaining part of each of the forces F₁ and F₂ issupported through the bearing 10 at the high pressure end of each rotor(bearing 10 of the female rotor 18 not shown in the drawings).

As a result of the geometry of the rotors defined by the toothingthereof the radial force F₁ is in most cases less than the radial forceF₂. Therefore there is a difference in length between the bearingbracket 11 and/or recess 13 of the male rotor 6 and the length of thebearing bracket 20 and/or recess 22 of the female rotor 18. This isindicated in FIG. 3 by distance "1".

As a result of pressurized oil being fed into the axial chambers 17, 33at the low pressure end of the rotors 6, 18 respectively, axial forcesF₇, F₈ (FIG. 3) are exerted on the rotors opposing the axial forces F₃and F₄ resulting from the compression of the gas. The axial forces F₇,F₈ compensate a part of the forces F₃ and F₄. The remaining part of theforces F₃ and F₄ is compensated through the bearings 10 of the rotors.

Due to the geometry of the rotors the axial force F₃ on the male rotor 6is as a rule larger than the axial force F₄ on the female rotor 18. Tocompensate this difference an additional axial force F₉ is exerted onthe male rotor 6.

According to the invention the longitudinal grooves 25 terminate at theend face of the bearing bracket to provide an open communication betweenthe grooves 25 and the space formed between the end face of the bearingbracket 11 and the bottom of the chamber 9 of the male rotor 6. As canbe seen in FIGS. 1-3 the passage of oil from this space towards therecess 13 is obstructed, whereby the oil pressure is maintained in thispart of the chamber 9. This results in the axial force F₉, which isexerted on the rotor 6. At the same time the axial force F₄ on thefemale rotor 18 will be smaller than the force F₃ and since the grooves24 on the bearing bracket 20 are not in open communication with thatpart of the chamber 19 no additional axial force is exerted on thefemale rotor. As the recess 22 terminates at the end face of the bearingbracket, the recess 22 is in open communication with the bottom part ofthe chamber 19, sot hat a built-up of oil pressure therein that isprevented.

The provision of bearing brackets at the low pressure ends of therotors, which brackets project into internal essentially cylindricalcavities provided in the rotors and extend over a significant part ofthe length of rotors, results in a bearing arrangement having anexcellent stiffness and capable of supporting high radial loads on therotors. In combination with the comparatively small distance between thebearings at opposite ends of each rotor the deflection of the rotorsresulting from the gas pressure is even further reduced. The bearingarrangement according to the invention is also capable of counteractingthe axial forces on the rotors without having to provide complexadditional thrust bearings.

The bearing arrangement of the rotary screw compressor according to theinvention permits a considerable increase of the radial and axial forcesover existing bearing arrangements, resulting in an increase of theallowable differential pressure and discharge pressure of the screwcompressor.

We claim:
 1. A rotary screw compressor comprising a casing defining aworking space having a high pressure end and a low pressure end, a malerotor and a female rotor cooperating with the male rotor, the male andfemale rotors being enclosed in the working space defined by the casing,the casing having an outlet port at the high pressure end of the workingspace, a discharge outlet connected to the outlet port at the highpressure end of the working space and a suction inlet at the lowpressure end of the working space, at least one rotor being rotatablysupported at an end thereof through a bearing arrangement comprising abearing bracket, means fixing the bearing bracket relative to thecasing, the bearing bracket having a substantially cylindrical outercircumferential surface, the bearing bracket projecting into an axialcavity provided in the rotor so as to form a first chamber between thebracket and the rotor, the bracket being provided with an oil feedchannel to feed oil into the first chamber, wherein at least one rotoris rotatably supported at the low pressure end thereof through thebearing arrangement, the corresponding bearing bracket being mounted atthe low pressure end of the working space and having an outercircumferential surface that is provided with at least a grooveconnected to the oil feed changed and a recess connected to an oildrainage channel provided in the bearing bracket, and sealing meansprovided between the first chamber and the working space of thecompressor.
 2. A rotary screw compressor according to claim 1, furthercomprising an end cover fixed to the bearing bracket and wherein the endface at the low pressure end of a rotor, the end cover, the casing, andthe corresponding bearing bracket define a second chamber, the secondchamber being connected to an oil feed channel.
 3. A rotary screwcompressor according to claim 1, wherein the outer circumferentialsurface of at least one of the bearing brackets is provided with twolongitudinal grooves and one recess, the recess being located on theside of the bearing bracket radially opposite the outlet port and beingconnected to the oil drainage channel, the longitudinal grooves beinglocated at either side of the recess and being connected to the oil feedchannel.
 4. A rotary screw compressor according to claim 1, wherein theouter circumferential surface of at least one of the bearing brackets isprovided with two longitudinal grooves and one recess, the recess beinglocated on the side of the bearing bracket radially opposite the outletport and being connected to the oil drainage channel, the longitudinalgrooves being located at either side of the recess and being connectedto the oil feed channel, and wherein the edges of the longitudinalgrooves adjacent the recess are situated in a common plane through theaxis of the bearing bracket at an equal distance from the recess, and inthat the edges of the longitudinal grooves most distant from the recessare each situated in a plane inclined at an angle α to the common plane.5. A rotary screw compressor according to claim 1, wherein each recesshas a maximum length of 0.7 times the length of the bearing bracket. 6.A rotary screw compressor according to claim 1, wherein the length ofthe bearing bracket of the male rotor and the length of the recessthereof is less than the length of the bearing bracket of the femalerotor and the recess thereof.
 7. A rotary screw compressor according toclaim 1, wherein a groove connected to the oil feed channel on thebearing bracket of the male rotor and a recess on the bearing bracket ofthe female rotor terminate at the end face of the corresponding bearingbracket, and in that each recess on the bearing bracket of the femalerotor are located spaced from the end face of the corresponding bearingbracket.
 8. A rotary screw compressor according to claim 1, wherein atleast one of the rotors is provided with a ring shoulder protruding fromits low pressure end, the sealing means being provided between the ringshoulder and the casing.
 9. A rotary screw compressor according to claim1, wherein at least one of the rotors is provided with sealing meansbetween the rotor and the corresponding bearing bracket.
 10. A rotaryscrew compressor according to claim 1, wherein a rolling contact bearingis provided between at least one of the rotors and the correspondingbearing bracket.
 11. A rotary screw compressor according to claim 1,wherein supply means are provided to supply oil to the oil feed channelsof the bearing brackets at a pressure approximately equal to thepressure of the gas to be compressed at the suction inlet, and in thatthe oil drainage channels of the bearing brackets are connected to anoil collector, the oil collector being connected to the supply means andbeing vented to the atmosphere.
 12. A rotary screw compressor accordingto claim 1, wherein supply means are provided to supply oil to the oilfeed channels of the bearing brackets at a pressure approximately equalto the pressure of the compressed gas at the discharge outlet, and inthat the oil drainage channels of the bearing brackets are connected toan oil collector, the oil collector being connected to the supply meansand to the suction inlet.
 13. A rotary screw compressor according toclaim 1, wherein the oil feed channels of the bearing brackets areconnected to an oil separator, the oil separator being connected to thedischarge outlet of the compressor, and in that the oil drainagechannels of the bearing brackets are connected to the suction inlet.